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Contract No. DE-AC36-08GO28308

Passive Two-Phase Cooling of

Automotive Power Electronics

Preprint

G. Moreno, J. R. Jeffers, S. Narumanchi, 

and K. Bennion

Presented at SEMI-THERM 2014 

San Jose, California

March 9–13, 2014

Conference Paper

NREL/CP-5400-61083

August 2014
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Passive Two-Phase Cooling of Automotive Power Electronics

Gilberto Moreno*, Jana R. Jeffers, Sreekant Narumanchi, and Kevin Bennion

National Renewable Energy Laboratory

Golden, CO, USA *

gilbert.moreno@nrel.gov

Abstract

Experiments were conducted to evaluate the use of a

passive two-phase cooling strategy as a means of cooling

automotive power electronics. The proposed cooling approach 

utilizes an indirect cooling configuration to alleviate some 

reliability concerns and to allow the use of conventional 

power modules. An inverter-scale proof-of-concept cooling 

system was fabricated and tested using the refrigerants 

hydrofluoroolefin HFO-1234yf and hydrofluorocarbon HFC245fa. Results demonstrated that the system can dissipate at 

least 3.5 kW of heat with 250 cm3 of HFC-245fa. An 

advanced evaporator concept that incorporates features to 

improve performance and reduce its size was designed. 

Simulation results indicate the concept’s thermal resistance 

can be 58% to 65% lower than automotive dual-side-cooled 

power modules. Tests were also conducted to measure the 

thermal performance of two air-cooled condensers—plain and 

rifled finned tube designs. The results combined with some 

analysis were then used to estimate the required condenser 

size per operating conditions and maximum allowable system 

(i.e., vapor and liquid) temperatures. 

Keywords

Power electronics, thermal management, two-phase heat 

transfer

Nomenclature

A area, cm2

P pressure, MPa

R”th unit/area-weighted thermal resistance, cm2

-K/W

T temperature, °C

Subscripts: 

a air

c critical (e.g., critical pressure)

htr heater

j junction

l liquid

r reference (e.g., reference area)

sat saturated

v vapor

1 Introduction

Improved thermal management is an enabling strategy to 

increase power densities, increase reliability, and reduce cost 

of electronic devices. Traditionally, single-phase liquid 

cooling strategies have been employed to cool power 

electronics in electric-drive vehicles. Improvements to the 

automotive power module’s thermal resistance, for the most 

part, have been focused on reductions to the passive stack 

thermal resistance. Significant improvements in power 

electronics thermal management can also be achieved using 

more aggressive cooling techniques (i.e., two-phase heat 

transfer). The high heat transfer rates of two-phase heat 

transfer allow for increased power density and specific power 

that may enable achieving the U.S. Department of Energy’s 

Advanced Power Electronics and Electrical Motors Program

technical targets. 

1.1 Conventional Automotive Power Electronics Cooling 

Strategies

Liquid cooling solutions are commonly used for cooling 

power electronic devices. Water-ethylene glycol (WEG)-

based cooling strategies are well understood in terms of 

performance and reliability and thus are widely used in 

automotive power electronics cooling applications. Liang [1] 

has experimentally characterized junction-to-liquid thermal 

resistances for the 2010 Nissan Leaf and 2010 Toyota Prius 

power modules. One of the more effective WEG-based 

cooling strategies is a dual-side-cooling approach 

implemented in the 2008 Lexus Hybrid [2, 3]. The dual-side 

cooling approach of this system allows for greater heat 

dissipation per die footprint as compared to the more 

traditional single-side cooled modules that are more typical of 

automotive applications. 

1.2 Two-Phase Based Cooling Alternatives

Two-phase-based cooling strategies have been studied in 

as a means of increasing device power densities. Ayers et al. 

[4] proposed a two-phase-based cooling strategy where the 

power electronics cooling system can be combined with the 

automotive air-conditioning system. In their approach, the 

power electronics were immersed in HFC-134a resulting in a 

reported 50% inverter volume reduction. Barnes and Tuma [5] 

evaluated passive and immersion two-phase cooling of power 

electronics. Heat fluxes in excess of 1,000 W/cm2 were 

reported using a passive two-phase cooling approach. Based 

on their tests, it was estimated that refrigerant requirements 

for a passive immersion cooling system could be on the order 

of 100 cm3 of refrigerant per kilowatt of heat dissipated. In 

addition, an in-situ degassing technique to remove air that 

may infiltrate the system was described. Performance 

characteristics of various air-cooled condenser designs 

operating in thermosyphon configurations and charged with 

hydrofluoroether fluids are reported by Tuma et al. [6]. 

Various two-phase-based thermosyphon cooling concepts 

have been explored for a variety of electronics cooling 

applications. Franco and Filippeschi [7] published a literature 

review of thermosyphons dissipating less than 1 kW of heat. 

They discuss differences in experimental design and attempt 
2 

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to explain the lack of generalized results for all data sources

reviewed. Difficulty in comparing data arises due to 

differences in physical design and possible presence of noncondensable gases in the system. Khodabandeh [8] explored 

the effects of fluid properties, system pressure, heat flux, mass 

flow rate, vapor friction, diameter of evaporator channel, and

tubing distance between the evaporator and condenser on 

thermosyphon performance. Isobutane was used as the 

working fluid for all experiments. Of all the parameters tested, 

heat flux and reduced pressure were the most influential. 

Results indicated that nucleate boiling was the dominant mode 

of heat transfer in the evaporator. Similarly, Ramaswamy et 

al. [9] investigated the combined effects of sub-cooling and 

system pressure on thermosyphons with fluorocarbon FC-72 

as the working fluid. It was determined that the combined 

effect of pressure and sub-cooling was to lower wall 

superheat. 

Agostini et al. [10] designed a small thermosyphon for use 

in electronics cooling. A mathematical model was used to 

design the condenser, and the modeling results were 

compared against experimental results. The system consisted 

of an evaporator mounted to the power electronics, a vapor 

riser, an air-cooled condenser, and a liquid downcomer. In 

their study critical heat flux was not reached but heat rejection 

of up to 1.5 kW was reported. Khrustalev [11] evaluated the 

performance of two evaporators used in parallel (two separate 

evaporators connected via manifold to the same condenser) in 

an electronics cooling thermosyphon application. Methanol 

and ethyl alcohol were chosen as working fluids. It was 

reported that a loop thermosyphon is advantageous over a 

traditional heat pipe for electronics cooling. In a similar 

application, Pal et al. [12] used a small thermosyphon to cool 

a computer microprocessor with water and performance fluid 

PF5060. The evaporator and condenser were separated by a 

rising vapor tube and a liquid downcomer tube. Water was 

observed to outperform the dielectric liquid, providing a

thermal resistance about 40% that of PF5060. In addition to 

studying the thermal performance of the two liquids, Pal et al. 

[12] studied the effects of inclination and found that 

evaporator surface temperatures remain uniform for 

inclination angles from -58° to +36°. 

In this study, we evaluate a passive two-phase strategy as 

a means of improving thermal management of automotive 

power electronics. The intent is to utilize the high heat 

transfer rates associated with phase-change heat transfer to 

increase device power densities. A proof-of-concept cooling 

system, designed to cool six power modules, was fabricated. 

Experiments were conducted with refrigerants HFO-1234yf 

and HFC-245fa to determine if the system could dissipate the 

expected automotive heat load. Simulations were then 

performed to estimate the performance of an advanced 

evaporator design. The simulation thermal performance 

results were then compared with the performance of 

conventional automotive WEG based cooling systems. 

Condenser experimental results and analysis lead to condenser 

sizing requirements based on operating conditions and a 

maximum allowable system temperature. 

2 Experimental Apparatus and Procedures

A proof-of-concept passive two-phase cooling system was 

designed and fabricated. Figure 1 shows the schematic of the 

cooling system that consists of an evaporator and an aircooled condenser. The design cools automotive power 

modules consisting of six Delphi discrete power 

modules/switches. The six power modules represent a voltage 

source inverter system required to power an automotive 

electric machine/motor. 

Figure 1: Schematic of the passive two-phase cooling system. 

The system was designed with a maximum operating 

pressure of 1.03 MPa (150 psi). Structural finite element 

analyses (FEA) were used to design the system components 

for the elevated pressure requirements. Once fabricated, the 

system was hydrostatically pressure-tested to verify the 

system’s pressure rating. The system’s pressure rating 

provided flexibility by allowing it to be charged with 

refrigerant HFO-1234yf or HFC-245fa. System vapor 

pressure was measured using an absolute pressure transducer.

System temperatures were measured using calibrated K-type 

thermocouples for the inlet-air, outlet-air, liquid, vapor, and 

heater temperatures. 

A cross-sectional view of the evaporator is provided in 

Figure 2. The evaporator consists of a stainless steel cylinder 

and three copper cold plates. For this proof-of-concept 

system, steel was used instead of aluminum because it is 

easier to weld leak-free joints with steel. As will be discussed 

25.4 cm (10 in)

20.5 cm 

(8.1 in)

4.5 cm 

(1.75 in)

25.4 cm (10 in)

 Ø 6.98 cm 

(2.75 in) Tl

T v,

P v

Ta, inlet

Ta, outlet 

Thtr
3 

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later, future designs will incorporate brazed aluminum 

evaporators. The copper cold plates were machined with 4-

mm-tall fins, each of which has a wetted surface area of about 

55 cm2

. Because two-phase cooling provides high heat 

transfer rates, the fins are not expected to have a significant 

effect on the evaporator thermal resistance. However, the fins 

were included as a means of increasing surface area in an 

attempt to delay dry-out. No surface enhancement coatings 

were applied to the cold plates, leaving the finned surfaces 

smooth.

Figure 2: Cross-sectional view of the evaporator vessel.

The evaporator was designed to cool six power modules, 

two per cold plate. For the tests, six Watlow ceramic heaters 

were used in place of the power modules. The dimensions of 

the ceramic heaters (25 mm × 15 mm × 2.5 mm) were similar 

to those of the Delphi modules. Each heater generates up to

about 580 W of heat (total power for six heaters: 3.5 kW). The 

ceramic heaters were externally attached to the evaporator 

using thermally conductive grease as the thermal interface 

material. The temperatures of the heaters were measured via 

thermocouples embedded within the ceramic heaters. The cold 

plates with attached heaters were then bolted onto the 

evaporator vessel. O-rings seals between the vessel and the 

cold plates prevented refrigerant leakage. The interchangeable 

cold plate design of the evaporator allows for testing of 

various two-phase-based cooling technologies for power 

electronics and other high-power density applications. The 

evaporator’s flexible design also allows testing of different 

boiling/evaporation enhancement techniques (e.g., boiling 

enhancement coatings and finned surfaces). 

The condenser consisted of a finned-tube, air-cooled heat 

exchanger. The fins were louvered and fabricated from 0.15 

mm thick aluminum sheets (Figure 3). The total air-side 

surface area was calculated to be approximately 29,000 cm2

. 

Twenty 0.95-cm (3/8-in)-outer-diameter copper tubes were 

arranged in two rows (see Figure 1). Two tubes, one from 

each row, were connected at the top via U-bend fittings. A 

manifold header connected all tubes at the lower end. Two 

condensers were fabricated; one with plain tubes and one with 

rifled tubes. The rifled features, shown in Figure 3, were 

evaluated as a means of enhancing condensation heat transfer. 

Per the manufacturer, the condensation-side surface areas 

were 1,190 cm2 and 1,250 cm2 for the plain and rifled 

condensers, respectively. Heat from the system was rejected 

to air by means of a 17.8-cm (7-in) diameter automotive axial 

fan mounted to the condenser. The fan operates on 12 V and 

consumes approximately 38 W of power, calculated by 

voltage and current measurements. Two 2.54-cm-innerdiameter tubes connected the condenser manifold to the 

evaporator. The tubes were sized large enough to prevent 

system pressure oscillations associated with liquid 

entrainment by the rising vapor, within the heat load range 

tested. 

Figure 3: Images of the finned-tube condenser (top-left), 

rifled tube (top-right), and louvered-fin design (bottom).

Procedures were established to charge the system with 

refrigerant. The system was charged using a transfer tube that 

contained 250 cm3 of saturated and oil-free refrigerant. This 

volume translated to approximately 330 grams of HFC-245fa 

or 280 grams of HFO-1234yf. Prior to charging the system 

with refrigerant, the air in the system was removed using a 

vacuum pump. Pressure within the system was allowed to 

decrease to about 2 Pa to indicate that most of the air was

evacuated from the system. A valve between the transfer tube 

and the system was then opened, allowing 250 cm3 of 

refrigerant to drain into the system. Once the refrigerant was 

transferred, the valve between the transfer tube and the system 

was closed, and the transfer tube was disconnected from the 

system. Measurements of the vapor temperature and pressure 

confirmed saturated conditions verifying that no air was 

present within the system. 

Cold plates
4 

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Thermal characterization experiments were initiated after 

saturated conditions were verified. First, the fan was powered 

to pull ambient air (Ta inlet = 25°C) through the condenser. 

The six heaters were then powered using two Agilent direct 

current power supplies. Heat loads from the heaters ranged 

from 250 W to 3,500 W. The system was allowed to reach 

equilibrium conditions (about 15 minutes) at each power level 

before increasing the power. Measurements of the system and 

heater temperatures combined with heat load measurements 

were used to compute the evaporator and condenser thermal 

resistances at various power levels. 

3 Uncertainty Analysis

An analysis was conducted to quantify the uncertainty in 

the measured experimental values according to the procedures 

outlined by Dieck [13]. The procedure consisted of gathering 

systematic and random uncertainties in all measured variables. 

The propagation-of-error equation was then used to estimate 

the uncertainties in the calculated values. The uncertainty for 

the thermal resistance values is estimated to be approximately 

±9%. After calibration, the uncertainty in a thermocouple was 

conservatively estimated at ±0.1°C, while the uncertainties in 

the heat measurements were estimated to be ±1%. All stated 

uncertainties were calculated at a 95% confidence level.

4 Results and Discussion

Experiments were conducted to measure the thermal 

performance of the cooling system with refrigerants HFC245fa (330 grams) and HFO-1234yf (280 grams). The unit

(area-weighted) thermal resistance of the evaporator is plotted 

versus the heat dissipated in Figure 4. The heater-to-liquid 

resistance is defined per Equation 1.

 R”th, evaporator = (Thtr ����� − Tl)

Total power

× Ar (1)

where Thtr ����� is the average temperature of the six ceramic 

heaters as measured by the thermocouples embedded within 

each heater, and Tl is the refrigerant liquid temperature. The 

total power is the total heat dissipated by the system, and Ar is 

the total surface area of the six heaters (22.5 cm2

). 

Tests were performed with increasing and decreasing heat 

fluxes to evaluate for hysteresis effects. Results showed that 

using an increasing or decreasing heat rate had little effect on 

the thermal performance of the system. 

With 250 cm3 of HFC-245fa, the system was capable of 

dissipating 3.5 kW of heat under steady-state conditions 

without reaching dry-out (i.e., critical heat flux). The 

refrigerant-volume-to-heat-dissipated ratio, at maximum heat 

load, was 71 cm3

/kW, which is somewhat less than the 100 

cm3

/kW estimate by Barnes and Tuma [5]. At 3.5 kW, the 

heaters are at their maximum allowable power rating, thus it 

was not possible to test to higher heat loads. However, results 

indicate the system may dissipate even greater heat loads. The 

3.5 kW of heat dissipation is noteworthy because it is a 

conservative estimate on the inverter heat dissipation 

requirement for a 55-kW electric traction-drive system. With 

HFO-1234yf, tests were limited to lower heat loads due to 

HFO-1234yf’s higher operating pressures. At 1.25 kW of heat 

dissipation, HFO-1234yf’s saturated temperature reached

42°C, which corresponds to a pressure of about 1.1 MPa—the 

maximum operating pressure of the system. Increasing the 

heat dissipation with HFO-1234yf requires increasing the 

condensing and/or pressure capacity of the system.

Figure 4: Averaged evaporator (heater-to-liquid) thermal 

resistance values versus the heat dissipated. The error bars 

indicate maximum and minimum values.

Because ceramic heaters attached via thermal interface 

material were used as a substitute for actual power modules 

attached via bonded interface materials (e.g., thermoplastics, 

solder), the thermal resistance values are not necessarily 

indicative of the thermal performance of the cooling system. 

Even so, it was still possible to draw conclusions from the 

thermal resistance values. Most notably, compared at the same 

heat loads, HFO-1234yf produced thermal resistance values 

that were 22% - 47% lower than those produced with HFC245fa. HFO-1234yf’s lower thermal resistance values are 

likely associated with a quicker onset of boiling at a lower 

wall superheat (at lower heat loads) and higher heat transfer 

rates. At a 1 kW heat dissipation, the system temperatures 

were approximately the same with both refrigerants (i.e., Tsat

= 40°C in both cases). A Tsat of 40°C corresponds to reduced 

pressures of 0.07 and 0.31 for HFC-245fa and HFO-1234yf, 

respectively. HFO-1234yf’s higher reduced pressure 

contributes to both its higher boiling heat transfer rates and 

lower boiling incipient superheat. A prior study [14], has 

shown that for boiling on plain copper surfaces, HFO-1234yf 

yields higher heat transfer rates as compared with HFC-245fa 

when evaluated at the same saturation temperature. 

As shown in Figure 4, the heater-to-evaporator resistances 

for both refrigerants decrease with increasing heat dissipation. 

This effect is associated with an increased contribution from 

two-phase heat transfer (i.e., boiling/evaporation) at higher 

0.00

0.25

0.50

0.75

1.00

0 1,000 2,000 3,000 4,000

R"th (cm2-K / W)

Heat Dissipated (W)

 HFC-245fa

 HFO-1234yf

HFO-1234yf: Tsat = 42°C

Psat = 1.1 MPa (154 psia)

HFC-245fa: Tsat = 75°C

Psat = 0.7 MPa (106 psia)
5 

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power levels, which improves thermal performance. The 

system’s temperature and pressure increased with increasing 

heat dissipation. Boiling heat transfer rates will increase with 

increasing pressure; thus, this effect also contributes to the 

observed heater-to-evaporator decreasing resistance trend. 

The performance results provided in Figure 4 were 

obtained using an evaporator without enhanced surfaces. The 

use of boiling enhancement coatings such as 3M’s copper 

microporous coating [15] within the evaporator will enable 

both increased heat dissipation and lower thermal resistance 

values. The improved performance of an advanced evaporator 

design employing these enhancement techniques will be 

discussed later.

The unit-thermal resistances for the condensers are 

provided in Table 1 for both refrigerants. The vapor-to-air 

resistance includes the condensation-side and the air-side 

resistances and is defined per Equation 2.

R”th, condenser = (Tv − T�

a)

Total power

× Ar (2)

where Tv is the refrigerant vapor temperature and T�

a is the 

average inlet air temperature as measured by two 

thermocouples just upstream of the condenser. Total power is 

the total heat dissipated and Ar is the frontal area (i.e., 

footprint) of the condenser. The stated condenser 

performances are for a passive condenser design cooled via an 

automotive axial fan. The total fan parasitic power was 

measured to be 38 W. An estimate of the air flow rate was 

calculated by imposing an energy balance on the air-side (i.e., 

inlet and outlet to the condenser) and the total heat dissipated. 

Assuming steady- state conditions and incorporating the 

appropriate air thermal properties, the estimated air 

volumetric flow rate is 0.12 m3

/s (250 ft3 per minute).

Condensation-side enhancements from the rifled tubes 

were found to reduce the overall condenser thermal 

resistances by 18% and 25% for HFC-245fa and HFO-1234yf, 

respectively. The results show that the rifled structures are 

more effective at enhancing condensation heat transfer with 

HFO-1234yf. Moreover, better performance is achieved with 

HFO-1234yf. For the plain tube condenser, HFO-1234yf 

yielded thermal resistance values that were about 13% lower 

than those with HFC-245fa. For the rifled tube condenser, 

HFO-1234yf yielded thermal resistance values that were

about 20% lower than those with HFC-245fa. 

R”th (cm2 - K/ W)

Plain Rifled

HFC-245fa 9.30 7.58

HFO-1234yf 8.12 6.06

Table 1: Condenser thermal resistance values calculated 

using Equation 2.

4.1 Advanced Evaporator Concept

The previously discussed work demonstrated the 

feasibility and performance of a passive two-phase cooling 

system. Knowledge gained from that work was then used to 

develop a more compact/lightweight evaporator with 

improved thermal performance. This section describes the 

research work related to the analysis of the advanced 

evaporator design and the analyses work conducted to size the 

system condenser.

Figure 5: Simple schematic of the advanced evaporator 

designed to cool six power modules.

Techniques to improve the evaporator performance and 

reduce its size have been identified. The techniques 

incorporate features to reduce the evaporator thermal 

resistance while utilizing low-cost fabrication techniques and 

materials (e.g., aluminum). The concept uses an indirect 

cooling approach to be compatible with conventional power 

electronic packages (i.e., silicon on ceramic substrate). 

A simple schematic of the advanced evaporator design is 

shown in Figure 5. To protect potentially sensitive intellectual 

property, certain evaporator features were omitted from the 

figure. The evaporator was designed to cool six Delphi 

discrete power switches. Although the initial tests with the 

proof-of-concept system were conducted using 250 cm3 of 

refrigerant charge, the advanced design will utilize less 

refrigerant—about 180 cm3 (HFC-245fa = 240 grams, HFO1234yf = 200 grams). To put these quantities into perspective, 

a 2010 Toyota Camry requires 510 grams of refrigerant 

(HFC-134a) for the cabin’s air-conditioning system [16]. 

Reducing the refrigerant quantity is an effort to reduce the 

system cost, weight, and size. The system design combined 

with the system refrigerant volume should allow for operation 

at various degrees of inclination. Future tests are planned to 

evaluate the effect of orientation on system performance. 

The thermal performance of the advanced evaporator 

design was simulated through FEA. A computer-aided-design 

model of the Delphi power modules bonded to the advanced 

evaporator design was first generated. The computer-aideddesign model incorporated all the thermal resistance interfaces 

within the package stack including the solder layers within the 

package and a thermoplastic-type bonded interface between 

the module and the evaporator base plate. The model was then 

imported into ANSYS Workbench for thermal analysis. The 

FEA imposed heat transfer coefficient boundary conditions to 

simulate boiling heat transfer from enhanced surfaces within 

the evaporator. In a prior studies [14, 17], the boiling heat 

transfer coefficients for HFC-245fa and HFO-1234yf were 

2.54 cm 

(1 in)

7 cm (2.75 in)

Length: 

17.8 cm (7 in) 

Power 

devices
6 

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experimentally measured with enhanced surfaces (3M 

microporous boiling enhancement coating) at various 

saturated temperatures. In those studies, the measured heat 

transfer coefficient values exceeded 100,000 W/m2

-K within a 

wide heat flux range. Moreover, the performance of the two 

refrigerants was found to be similar for boiling on 

microporous coated surfaces. Because uniform boiling may 

not occur with the evaporator surfaces, a conservative 

estimate on the heat transfer coefficients was imposed (50,000 

W/m2

-K) for these simulations. 

The unit/area-weighted thermal resistance results 

(junction-to-liquid) as predicted by FEA are provided in Table

2. The insulated gate bipolar transistor’s area and maximum 

temperature (i.e., junction temperature) combined with the 

refrigerant temperature were used to calculate the thermal 

resistance values. Two evaporator designs were analyzed— 

aluminum-based and copper-based. The aluminum fabrication 

is the more practical design while the copper fabrication is 

more of an idealistic concept aimed at reducing the package 

stack resistance. For comparison, the thermal resistance 

(junction-to-liquid) of the 2008 Lexus Hybrid dual side and 

liquid (WEG)-cooled power module is also shown in Table 2. 

The thermal resistance value of 0.33 cm2

-K/W listed for the 

dual-side-cooled module was calculated using performance 

data from Sakai et al. [3] and the module’s insulated gate 

bipolar transistor footprint as measured by Buress et al. [2]. 

R”th (cm2 - K/ W)

Two-phase: aluminum evaporator 0.14

Two-phase: copper evaporator 0.12

2008 Lexus Hybrid, [3] 0.33

Table 2: The FEA estimated evaporator (junction-to-liquid) 

thermal resistances for aluminum and copper evaporators. 

Performance data for an automotive cooling system is 

provided for comparison. 

The FEA predicted thermal resistance values of the 

aluminum (i.e., 0.14 cm2

-K/W) and copper (i.e., 0.12 cm2-

K/W) evaporator modules are about 58% and 65% lower, 

respectively, than that of the dual-side, WEG-cooled 2008 

Lexus system. That proposed two-phase cooling configuration 

outperforms the dual-side-cooled (WEG coolant) package 

because of improved heat transfer from two-phase cooling and 

also because of differences in the passive stack within the 

packages. Future work is planned to experimentally measure 

the junction-to-liquid thermal resistance for the advanced 

evaporator design. In addition to the thermal improvements, 

the passive cooling approach also improves system efficiency.

An immersion cooling (two-phase) strategy may allow for 

even greater thermal enhancements. However, an indirect 

cooling approach was used in this case to allow for its use with 

more traditional power modules and to alleviate some reliability 

concerns associated with immersing electronics in a refrigerant. 

4.2 Condenser Sizing Implications

An analysis was conducted to estimate condenser size 

based on operating conditions, a maximum allowable cooling 

system temperature (i.e., vapor and liquid), and the 

experimentally measured condenser thermal resistance values 

provided in Table 1. The operating conditions used for this 

analysis were: 3.3 kW of steady-state heat dissipation

(estimated requirements for a 55 kW traction drive inverter) 

and 43°C inlet air temperature. The estimated condenser 

frontal area (i.e., frontal footprint) requirements are plotted 

versus the system temperature in Figure 6. In the figure, the 

lower system temperature of 65°C, used for this analysis, 

corresponds to the coolant (i.e., WEG) temperature typical of 

existing automotive power electronics systems. The solid line 

curves for both refrigerants are the sizing requirements for the 

finned-tube condenser per the performance data reported in 

Table 1. The Tj curve in Figure 6 represents the maximum 

junction temperature and was calculated using the thermal 

resistance data for an aluminum evaporator and a six power 

module/switch system configuration. 

Figure 6: Condenser sizing requirements at various system 

temperatures. Barnes and Tuma data from [5].

Because typical automotive condensers are constructed 

using a brazed, folded-fin design as shown in Figure 6, a 

simplified analysis was conducted to estimate the sizing 

requirements for a folded-fin type condenser. For this 

analysis, the ratios of the air-side surface area to frontal 

surface area were measured for the finned-tube and folded-fin 

designs shown in Figure 7. The air-side-to-frontal-surfacearea ratios were calculated to be 55.5 mm2

/mm2 for the 

finned-tube and 86.7 mm2

/mm2 for folded-fin condensers, 

assuming a condenser thickness of 4.5 cm for both cases. The 

folded-fin frontal area requirements were estimated by 

matching the air-side surface area for the folded-fin design to 

those of the finned-tube design at the various system 

temperatures. The folded-fin frontal area requirements were 

then calculated via the surface area ratio (86.7 mm2

/mm2) and 

the results are shown Figure 6. This simplified analysis 

assumes that the air-side is the dominant thermal resistance, 

and air velocities are similar to those provided by the fan used 

0

25

50

75

100

125

150

175

200

0

500

1,000

1,500

2,000

50 60 70 80 90 100

Tj (°C)

Frontal Area (cm2)

System Liquid/Vapor Temperature (°C)

Tj

finned tube

folded-fin

HFC-245fa

HFO-1234yf

Barnes and Tuma (2009)
7 

This report is available at no cost from the National Renewable Energy Laboratory (NREL) at www.nrel.gov/publications.

in this study. For reference, the performance of a folded-fin 

condenser operating in a thermosyphon configuration, as 

reported by Barnes and Tuma [5], is provided in Figure 6. In 

that study, a condenser thermal resistance of 4.6 cm2

-K/W is 

reported using a hydrofluoroether fluid and an air velocity and 

pressure drop of 2.2 m/s and 44 Pa, respectively. The 

estimated condenser sizing requirements, per the performance 

reported by [5], are similar to the results predicted for HFC245fa. 

The size requirements of the condenser can then be 

estimated using Figure 6 and a maximum allowable system 

temperature. For low-temperature applications typical of 

existing automotive power electronics coolant temperatures 

and conventional silicon devices, the use of HFO-1234yf is 

recommended because it provides better performance. At a 

system temperature of 75°C and with HFO-1234yf, the 

condenser frontal size requirements are estimated to be about 

400 cm2 (20 cm × 20 cm) with a maximum T j of about 125°C. 

HFC-134a may also be considered in this case because it is a 

refrigerant currently used in automotive systems and because 

prior studies [17-20] have shown that its heat transfer 

performance (boiling and condensation) is similar, if not 

slightly better than, that of HFO-1234yf. 

Figure 7: Images of the finned-tube condenser (top) and a 

brazed folded-fin condenser (bottom). 

Operating at higher system temperatures has advantages 

because it allows for a more compact condenser. However, 

higher temperature operation may be dependent on the 

development of higher temperature auxiliary electronic 

devices (e.g., capacitors) that are packaged within the inverter. 

With this in mind, if higher system temperatures are practical, 

HFC-245fa is recommended due to its higher critical 

temperature (TC = 154°C) and lower operating pressures. At a 

system temperature of 95°C and a T j = 148°C, the condenser 

size requirements would decrease to about 310 cm2 (17.6 cm × 

17.6 cm). 

5 Conclusions

A proof-of-concept inverter-scale passive two-phase 

cooling system was designed and fabricated. Experiments 

demonstrated that the system can dissipate at least 3.5 kW of 

heat with 250 cm3 of HFC-245fa and total fan parasitic power 

consumption of 38 W. 

The performance of the finned-tube condensers (plain and 

rifled) operating in the passive (thermosyphon) configuration 

was experimentally measured. The rifled tubes were found to 

decrease the thermal resistance for HFC-245fa and HFO1234yf by 18% and 25%, respectively. Additionally, 

condenser thermal resistances were measured to be 13% to 

20% lower with HFO-1234yf as compared with HFC-245fa. 

An advanced evaporator design concept was conceived 

that improves thermal performance and reduces the 

evaporator size. The design utilized an indirect cooling 

configuration and thus can be used with conventional power 

module packages. FEA indicates the evaporator design can 

decrease the junction-to-liquid resistance by 58% and 65% as 

compared to the dual-side-cooled automotive power module. 

Experimental results and analyses were used to size the 

system condenser based on estimated inverter operating 

conditions at various maximum allowable system 

temperatures. For low system temperature applications, HFO1234yf is recommended, and the condenser frontal area 

requirements are estimated to be 400 cm2

. For higher system 

temperature operation, HFC-245fa is recommended due to its 

higher critical temperature. At a system temperature of 95°C, 

the condenser size can be decreased to 310 cm2

. However, 

higher temperature operation would likely be contingent on 

the development of higher temperature auxiliary power 

electronic devices (e.g., capacitor). 

Acknowledgments

The authors would like to thank Susan Rogers and Steven 

Boyd, Technology Development Managers of the U.S. 

Department of Energy Vehicle Technologies Office, 

Advanced Power Electronics and Electric Motors Program, 

for supporting this work. They also wish to thank Charlie 

King (National Renewable Energy Laboratory) for helping to 

fabricate the experimental system and Mary Koban (DuPont) 

for supplying the HFO-1234yf refrigerant.

References

[1] Z. Liang, "Power Device Packaging," Oak Ridge 

National Laboratory, 2012 U.S. DOE Hydrogen and Fuel 

Cells Program and Vehicle Technologies Program 

Annual Merit Review, 2012.

[2] T. A. Burress, C. L. Coomer, S. L. Campbell, A. A. 

Wereszczak, J. P. Cunningham, L. D. Marlino, L.E. 

Seiber, and H.T. Lin, "Evaluation of the 2008 Lexus LS 

600H Hybrid Synergy Drive System," Oak Ridge 

National Laboratory ORNL/TM-2008/185, 2009.

2.54 cm

2.54 cm
8 

This report is available at no cost from the National Renewable Energy Laboratory (NREL) at www.nrel.gov/publications.

[3] Y. Sakai, H. Ishiyama, and T. Kikuchi, "Power Control 

Unit for High Power Hybrid System," SAE Power, vol. 1, 

2007.

[4] C. W. Ayers, J. C. Conklin, J. S. Hsu, and K. T. Lowe, 

"A Unique Approach to Power Electronics and Motor 

Cooling in a Hybrid Electric Vehicle Environment," in 

2007 Vehicle Power and Propulsion Conference, IEEE, 

2007, pp. 102-106.

[5] C. M. Barnes and P. E. Tuma, "Practical Considerations 

Relating to Immersion Cooling of Power Electronics in 

Traction Systems," IEEE Transactions on Power 

Electronics, vol. 25, pp. 2478-2485, 2010.

[6] P. E. Tuma, B. O. Fayemi, and L. J. Stang, "Condenser 

Design for Thermosyphons Utilizing Segregated 

Hydrofluoroether Working Fluids," in Semiconductor 

Thermal Measurement and Management Symposium, 

SEMI-THERM 2007, 2007, pp. 162-167.

[7] A. Franco and S. Filippeschi, "Closed Loop Two-Phase 

Thermosyphon of Small Dimensions: a Review of the 

Experimental Results," Microgravity Science and 

Technology, vol. 24, pp. 165-179, 2011.

[8] R. Khodabandeh, "Heat transfer in the Evaporator of an 

Advanced Two-Phase Thermosyphon Loop," 

International Journal of Refrigeration, vol. 28, pp. 190-

202, 2005.

[9] C. Ramaswamy, Y. K. Joshi, W. Nakayama, and W. B. 

Johnson, "Combined effects of sub-cooling and operating 

pressure on the performance of a two-chamber 

thermosyphon," IEEE Transactions on Components and 

Packaging Technologies, vol. 23, pp. 61-69, 2000.

[10]F. Agostini, T. Gradinger, and C. de Falco, "Simulation 

Aided Design of a Two-Phase Thermosyphon for Power 

Electronics Cooling," IECON 2011-37th Annual 

Conference on IEEE Industrial Electronics Society, pp. 

1560-1565, 2011.

[11]D. Khrustalev, "Loop Thermosyphons for Cooling of 

Electronics," in 18th IEEE Semi-Therm Symposium, 

2002, pp. 145-150.

[12]A. Pal, Y. K. Joshi, M. H. Beitelmal, C. D. Patel, and T. 

M. Wenger, "Design and Performance Evaluation of a 

Compact Thermosyphon," IEEE Transactions on 

Components and Packaging Technologies, vol. 25, pp. 

601-607, 2002.

[13]R. H. Dieck, Measurement Uncertainty: Methods and 

Applications, 4 ed. Research Park Triangle Park, NC, 

USA: ISA, 2007.

[14]G. Moreno, J. R. Jeffers, and S. Narumanchi, "Effects of

Pressure and a Microporous Coating on HFC-245fa Pool 

Boiling Heat Transfer " in ASME Summer Heat Transfer 

Conference, Minneapolis, MN, USA, 2013, pp. 557-567.

[15] 3M, "3MTM Microporous Metallic Boiling Enhancement 

Coating (BEC) L-20227," ed: 3M, 2009.

[16] AutoZone. (12/2/13). Available: 

http://www.autozone.com

[17]G. Moreno, S. Narumanchi, and C. King, "Pool Boiling 

Heat Transfer Characteristics of HFO-1234yf on Plain 

and Microporous-Enhanced Surfaces," ASME Journal of 

Heat Transfer, vol. 135, p. 111014, 2013.

[18]D. Del Col, D. Torresin, and A. Cavallini, "Heat transfer 

and pressure drop during condensation of the low GWP 

refrigerant R1234yf," International Journal of 

Refrigeration, vol. 33, pp. 1307-1318, 2010.

[19]J. E. Park, F. Vakili-Farahani, L. Consolini, and J. R. 

Thome, "Experimental study on condensation heat 

transfer in vertical minichannels for new refrigerant 

R1234ze(E) versus R134a and R236fa," Experimental 

Thermal and Fluid Science, vol. 35, pp. 442-454, 2011.

[20]K.-J. Park, D. G. Kang, and D. Jung, "Condensation heat 

transfer coefficients of R1234yf on plain, low fin, and 

Turbo-C tubes," International Journal of Refrigeration, 

vol. 34, pp. 317-321, 2011.